1. Field of the Invention
The present invention relates to extruded manifolds with multiple passages. More specifically, the invention relates to extruded manifolds with multiple passages and cross-counterflow heat exchangers incorporating such extruded manifolds, which are suitable for use as commercial or residential condensers or evaporators.
2. Related Art
Air-cooling (or heating) cross-counterflow heat exchangers are well-known. In real-world applications, due to space limitations in many cases, the heat exchangers cannot be made with a large frontal surface area. In order to have sufficient overall heat transfer area to meet design performance requirements, the heat exchanger core has to be increased by adding rows of heat exchanger modules. The multi-row heat exchanger thus becomes necessary in practice. In current parallel-flow heat exchanger technology, such multi-row heat exchangers comprise a plurality of stacked, assembled modules, each module comprising a pair of spaced headers or manifolds interconnected by a plurality of spaced, parallel, flattened heat exchanger tubes and heat exchanger fins interposed between the heat exchanger tubes.
The concept of the cross-counterflow heat exchanger can be realized in multi-row heat exchanger designs. Typically, the cross-counterflow heat exchanger is arranged so that heat-exchanging air flows in a direction perpendicular to the surface plane of the heat exchanger core, which comprises several heat exchanging tube rows. As shown in FIG. 17, an in-tube heat exchanging fluid F is introduced into the heat exchanger core 1 at one side, and the air A enters the heat exchanger core 1 from the opposite side. In each tube row, the two fluids, in-tube fluid F and air A, flow normal to each other, as in a typical crossflow heat exchanger. However, if the flows between each tube row are considered, it will be appreciated that the two fluids A and F flow parallel to each other but in opposite directions, as in a typical counterflow heat exchanger. Overall, the heat exchanger core 1 is therefore considered to have a cross-counterflow arrangement.
Examples of such heat exchangers are disclosed in U.S. Pat. Nos. 4,829,780 and Re. 35,502 (originally 5,157,944), both to Hughes et al.
U.S. Pat. No. 4,829, 780 to Hughes et al. discloses an evaporator which comprises a number of integrally assembled heat exchange modules, each of which comprises a pair of spaced apart headers 12, 14 interconnecting a series of flat hollow heat tubes 40 in a manner to attain a serpentine flow between the headers.
U.S. Pat. Nos. 5,157,944 and Re. 35,502 to Hughes et al. disclose an evaporator including adjacent inlet and outlet headers 10 and 12 and adjacent intermediate headers 14 and 16 spaced apart from headers 10 and 12. Two U-shaped tubes 18 and 19 at the ends of headers 14 and 16 establish communication between the interiors of tubes 18 and 19. A plurality of flattened tubes 20, arranged in two rows, extend between the inlet and outlet headers 10 and 12 at one end and intermediate headers 14 and 16 at the other end.
Most conventional parallel-flow heat exchanges consist of a single row of tubes. In particular, in a conventional parallel-flow heat exchanger, two spaced manifolds or headers are provided, with a plurality of flat tubes fixedly connected therebetween to provide a plurality of fluid flow paths. Corrugated fins are positioned between the tubes. Typically, as least one baffle is positioned in at least one of the manifolds to partition the manifold into at least first and second chambers and redirect the fluid flow path to the other manifold.
When such a heat exchanger is used as a condenser, compressed refrigerant gas from an external compressor is introduced via an inlet pipe into the first chamber of the first manifold, and is distributed so that a portion of the gas flows through each of the flat tubes which is disposed upstream of the baffle, and into one end of the second manifold. The refrigerant flows through the second manifold towards its other end, and is distributed so that a portion of the refrigerant flows through each of the tubes disposed downstream of the baffle, and into the second chamber of the first manifold. As the refrigerant gas flows sequentially through the tubes, heat from the refrigerant gas is exchanged with the atmospheric air flowing through the corrugated fins. The condensed, sub-cooled liquid refrigerant in the second cavity of the first manifold flows out of the second cavity through an outlet pipe connected thereto.
As the heat-exchanging air flows into a single row condenser core of the type described above, it has the ambient atmospheric temperature uniformly on the cross-sectional surface. If the heat-exchanging fluid in a tube is a zeotropic mixture, its phase-changing process is no longer at a constant temperature.
A zeotrope is a mixture fluid made up of two or more types of compounds. Its evaporating and condensing temperatures vary in phase-changing processes. For example, in evaporation, because there is no unique boiling point for each compound, the components in the mixture do not vaporize at rates proportionally to their composition in the liquid state. The more volatile component vaporizes faster and more than the heavier component. Therefore, the more volatile component comprises a higher proportion of the composition in the vapor phase than in the liquid phase.
FIG. 18 shows a two-component zeotropic mixture phase diagram for two compounds A and B, where the compound B is the more volatile of the two components. When a subcooled liquid having the composition shown at point a is heated, the temperature of the mixture rises until it starts boiling, or reaches the "bubble point" (the point at which bubbles begin to appear in the liquid). At the bubble point, the liquid phase composition is read from the point b.sub.L and its vapor phase composition is read from the point b.sub.V. During boiling, the more volatile component in the mixture is preferentially vaporized, to increase the composition of the heavier component in the liquid phase, so that the system saturation temperature rises. At the point c, the compositions in both the vapor phase (at c.sub.V) and the liquid phase (at c.sub.V) are no longer the original values. As the last drop of liquid vaporizes, the mixture reaches the dew point line. This is the temperature at which liquid begins to appear when the zeotropic vapor is cooled. At the dew point, the vapor-phase composition is d.sub.V, and the liquid-phase composition is d.sub.L. With more heating, the mixture becomes a superheated vapor from d to e. This superheated vapor has the same composition as point a. During evaporation, from the time the first bubble appears to the time the last droplet vaporizes, the system evaporating temperature (or saturation temperature) increases. This increase of the saturation temperature from the bubble point to the dew point is called the "temperature glide." A similar analysis for condensation can be made from the phase diagram of FIG. 18.
The practical effect of the temperature glide in a heat exchanger is that, as shown in FIG. 19, as a mixture flows through the heat exchanger core at constant pressure, the evaporating (or condensing) temperature rises (or drops) from E1 (or C1) at the inlet to E2 (or C2) at the outlet of the evaporator (or condenser). A constant evaporating or condensing temperature process, which exists in the single-component fluid, does not occur in a zeotropic mixture fluid.
Due to temperature glide effects, the temperature differential between air and in-tube fluid at the inlet may be much higher than at the outlet. The temperature differential profile on the entire heat transfer surface could then be highly non-uniform. Similar conclusions can be drawn if the heat exchanger core is used as a zeotropic mixture fluid evaporator or a single-phase fluid heat exchanger, for example, an air-glycol/water radiator, an air-air charge air cooler, and so on. According to presently-known heat exchanger design practice, at certain temperature variation ranges in hot fluid and cold fluid, the more uniform the profile of temperature differentials between hot and cold fluids on the entire heat exchanging surface area, the more efficient is the heat exchanger performance. Therefore, it is necessary to find a way to improve the temperature differential profile in the heat exchanger.
The counterflow arrangement is thermodynamically superior to any other flow arrangement. Ideally, it is the most efficient flow arrangement producing the highest temperature change in each fluid compared to any other two-fluid flow arrangements in an exchanger for a given amount of surface area and fluid flow rates. Thus, we (the present inventors) have introduced the counterflow design concept into current micro-channel heat exchanger design to produce the cross-counterflow concept. However, this cross-counterflow arrangement is practically feasible only for a heat exchanger with a thicker core. This cross-counterflow heat exchanger utilizes the temperature variations in both heat exchange fluids (if any) to improve the heat exchanger performance. For two-phase zeotropic mixture fluids and single-phase fluids, because their temperatures change through the entire heat transfer process, the cross-counterflow concept can reduce the non-uniformity of the temperature differential profile between hot and cold fluids in heat exchangers, and increase the overall heat exchange capacity at the same temperature variation ranges.
This problem is addressed in U.S. Pat. No. 5,174,373 to Shinmura, which discloses a heat exchanger in which the header pipes 11 and 12 are divided into at least two longitudinal chambers by at least one dividing wall which extends in the longitudinal direction. A plurality of flat tubes 13 extend between the header pipes 11 and 12, the flat tubes 13a being provided with slits 13a at their ends for receiving the peripheral surfaces of the dividing wall. The flat tubes have a plurality of fluid paths 9 formed by a plurality of longitudinal partitions 8. Baffles can be provided in the header pipes to change the flow path.
U.S. Pat. No. 5,203,407 to Nagasaka discloses a heat exchanger having spaced apart headers which redirect flow from groups of tubes back and forth between the headers, the headers having both longitudinal and transverse partitions which divide the headers into a plurality of longitudinal passages. In the embodiment of FIGS. 16 and 17, the header 40 comprises a tank 15 diametrically divided to form a pair of sub-passages 8 and 12, and an end plat 16 which cooperates with the tank 15 to form a main passage 34. The sub-passage 8 serves as a distributing chamber and the sub-passage 12 serves as a collecting chamber. The header can also be formed by extrusion as shown in FIG. 18 to form three passages.
U.S. Pat. No. 5,228,315 to Nagasaka et al. also discloses a heat exchanger with multi-passage headers. These headers can be extruded, with as many as five passages.
U.S. Pat. No. 31,444 to Cragg et al. discloses a steam boiler condenser having groups of parallel tubes mounted between a pair of headers which redirect flow back and forth between the headers.
U.S. Pat. No. 3,181,525 to McKann discloses a group of parallel tubes having manifolds on each end, the manifolds being provided with dividing walls for redirecting the flow back and forth between the manifolds.
U.S. Pat. No. 3,675,710 to Ristow discloses parallel groups of tubes mounted between headers 11 and 12, the headers 11 and 12 being provided with transverse partitions 18 for redirecting the heat exchange fluid back and forth between the headers. The headers 11 and 12 are also provided with longitudinally-extending condensate drain pipes 29 extending between holes in the partitions 18 for to drain condensate as it forms in the tubes.
U.S. Pat. No. 4,190,101 to Hartmann discloses a heat exchanger having parallel tubes between a pair of headers, one of which has a wall divider 21 for directing a portion of the total flow out of the tubes down to the other header where the flow is returned to the other set of tubes.
U.S. Pat. Nos. 5,086,835 and 5,176,200 to Shinmura disclose a heat exchanger which comprises a number of integrally assembled heat exchanger cores, each of which comprises a pair of spaced apart headers interconnecting a series of flat hollow heat tubes 13, 23 in a manner to attain a serpentine flow between the headers.
U.S. Pat. No. 5,186,248 to Halstead discloses a heat exchanger, e.g. a condenser, which includes a pair of spaced apart tanks, one of which is a unitary extrusion 30, 130 which forms a longitudinally-extending main tank 32, 132 and a longitudinally-extending outlet tank 34, 134; while the other has only a single return tank 42, 142 formed therein.
U.S. Pat. No. 5,348,081 to Halstead et al. discloses a condenser which comprises two layers assembled heat exchange modules, each of which comprises a pair of spaced apart headers 14, 16 interconnecting a series of flat hollow heat tubes 18 in a manner to attain a serpentine flow between the headers. The headers 14 can be connected by a cross-over pipe 40.
U.S. Pat. No. 5,400,853 to Wolters discloses a heat exchanger in which one of the manifolds 16 includes a return chamber 28 from which a return tube 30 extends the remainder of the length of the manifold.
U.S. Pat. No. 5,582,239 to Tsunoda et al. discloses a heat exchanger in which the first tank includes a first partition which divides it into at least two chambers and the second tank includes a second partition which divides it into one fewer chambers than the first tank. The partitions can extend both transversely and longitudinally.
None of the above-discussed prior art addresses the problem of undue size in heat exchangers such as those disclosed by Hughes et al. comprising more than three or four integrally assembled heat exchange modules; or how extruded and/or multiple passage manifolds such as those used in conventional parallel flow heat exchangers, can be applied to reducing the size of cross-counterflow heat exchangers.
Further, none of the above-discussed multi-row, cross-counterflow heat exchangers can eliminate air gaps between each heat exchanger row or module. The heat exchanger design disclosed in U.S. Pat. No. 5,174,373 to Shinmura has no air gap between rows, but the theory on which the design is based restricts the design to the two-row case. Through numerical analysis and experimental tests, we know that the air gap between rows can cause an additional pressure drop. The air gap also can trap solid particles and other material, which block the air flow paths and cannot easily be removed or cleaned out, and thereby reduce heat exchanger performance. In addition, the air gap increases the heat exchanger core thickness.
It is to the solution of these and other problems to which the present invention is directed.